Those skilled in the art of antifriction bearing engineering know that axial tapered roller bearings are used for rotational transmission of axial loads and to absorb radial tilting moments caused by shaft deflection. For example, they are used as pilot bearings in motor vehicle transmissions. Similar to the axial tapered roller bearing previously disclosed in JP 2000-110829, axial tapered roller bearings of this type substantially comprise an upper pressure disk and a lower pressure disk and a number of tapered rollers arranged between the pressure disks. The rollers are kept at uniform intervals from one another in the circumferential direction by a disk cage. In this case, the inner surface of the upper pressure disk is formed as a circulation path running radially obliquely outward for the tapered rollers and is provided with a circumferential outer rim which, with its annular surface facing the tapered rollers, is in radial supporting contact with the outer end faces of the tapered rollers. On the other hand, the inner surface of the lower pressure disk is formed as a flat mating circulation path for the tapered rollers and is provided exclusively to support axial load components.
Plano-convex contact geometry is standard in such an axial tapered roller bearing between the tapered roller end faces and the outer rim of the one pressure disk of the bearing. In this, the annular surface of the outer rim facing the tapered rollers is formed as a planar surface and the outer tapered roller end faces have the form of a sphere. This has proven to be the reason that under continuous load conditions, the originally convergent gap geometry between the tapered roller end faces and the outer rim is ground in the vertical direction into a constant spacing or into a constant lubrication gap height, as a result of the abrasive wear of the outer rim surface and the tapered roller end faces. However, this hampers the necessary hydrodynamic pressure development in the vertical direction which is a precondition for transmission of the force component at the outer rim of the axial tapered roller bearing, which results from the bearing loading, the height of the contact point path and from the outer rim and running track angle. Only the circumferential component of the hydrodynamically active relative speed of the tapered roller end faces makes it still possible to build a closed load bearing lubricating film in this case. However, a specific minimum rotational speed of the bearing is needed for this purpose. At the same time, the pressure or the peak pressure in the lubricating film decreases, as a result of the abrasive wear, which forms a greater ground contact surface between the tapered roller end faces and the outer rim. Furthermore, in the event of abrasive wear of the contact surface, the roughness peaks of the tapered roller end faces and outer rim surface meet each other, so that those peaks are bent over or sheared off, producing loose particles that contaminate the lubricant. The deformation energy or the energy resulting from shearing off the roughness peaks corresponds to the frictional energy which, as thermal energy, heats up at the lubricant and the entire bearing. The heat in the lubricant in turn causes lower viscosity of the lubricant, and this additionally reduces the lubricant gap height. Moreover, the chemical stability of the lubricant is impaired, subjecting the bearing to heat stress which is no longer acceptable, and its functioning is restricted. As the loading further increases, the load bearing lubricating film becomes thinner only builds up to an insufficient extent. The friction between the tapered roller end faces and the outer rim becomes higher until the frictional heat which is produced can no longer be dissipated, and the lubricating film finally breaks down. As a result, the tapered roller end faces rub in direct contact on the outer rim surface, with the consequence that the lubricant is burned and the bearing ultimately “seizes solid”.